Automatic transmission

ABSTRACT

An automatic transmission includes a first planetary gear set; a second planetary gear set; a third planetary gear set; an input shaft; an output shaft; and five friction elements. A first ring gear is constantly locked. A third carrier is connected with a second ring gear to define a rotating member. The input shaft is constantly connected with a second carrier, and the output shaft is constantly connected with the rotating member. The five friction elements are adapted to selectively connect among the carriers, ring gears and sun gears. Each of at least seven forward speed-ratios and one reverse speed-ratio is achieved by an engaged state of two friction elements selected from the five friction elements.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a Divisional of U.S. application Ser. No.12/271,115, filed Nov. 14, 2008, which is based upon and claims thebenefit of priority from prior Japanese Patent Application No.2007-309539, filed Nov. 29, 2007, the entire contents of which areincorporated herein by reference.

BACKGROUND OF THE INVENTION

The present invention relates to a step automatic transmission employedas a transmission for vehicle.

Japanese Patent Application Publication No. 2004-176765 or U.S. Pat. No.6,648,791 (FIGS. 23 to 26) discloses an automatic transmission adaptedto achieve forward seven speed-ratios by using three planetary gearsets. In the automatic transmission disclosed in the above JapanesePatent Application, the forward seven speed-ratios are obtained by usingsix friction elements and three single-pinion-type planetary gear sets.This single-pinion-type planetary gear set has an advantage in transferefficiency and gear noise and also an advantage in durability because ofthe nonnecessity for reducing a diameter of pinion gear. Moreoversimilarly, the automatic transmission disclosed in the above UnitedStates Patent achieves forward six speed-ratios to forward eightspeed-ratios by using five friction elements and threesingle-pinion-type planetary gear sets.

SUMMARY OF THE INVENTION

However, in the technique disclosed in the above Japanese PatentApplication, at least six friction elements are necessary to achieve theseven forward speed-ratios. Therefore, there is a problem that thenumber of friction elements is large so that an increase of the numberof components and an increase in axial length are incurred.

Since the number of friction elements provided for achieving the forwardsix to eight speed-ratios is five in the technique of the above UnitedStates Patent, there is an advantage that the number of frictionelements is small as compared with that of the above Japanese PatentApplication, so that the number of components can be reduced. However,in the automatic transmission shown in FIG. 23 of the above UnitedStates Patent, members passing on a radially outer side of a ring gearof center one of three planetary gear sets form a three-layeredstructure. Generally, in the automatic transmission, lubricating oil isreleased from a shaft-center side by means of centrifugal force and thenis collected into an oil pan provided in a lower portion of theautomatic transmission via respective parts requiring to be lubricated.In the case that the connecting members such as a drum member areprovided in a multilayer structure on the radially outer side ofplanetary gear set, the lubricating oil is easy to be retained insidethe above-mentioned members. Since many of these members rotate at thetime of vehicle running, there is a problem that respective frictionsare increased to worsen a fuel economy.

On the other hand, in the technique shown in FIG. 25 of the above UnitedStates Patent, members passing on the radially outer side of planetarygear set are provided in a two-layered form. Hence, lubricating oil isresistant to the retention (a disrupted flow) as compared to thetechnique of FIG. 23. However, a multi-shaft structure, concretelythree-layered structure at a maximum is formed on a radially inner sideof a sun gear of input-shaft-side one of three planetary gear sets.Hence, dimensions of the sun gear are restricted so that there is aproblem that a degree of freedom to design a gear ratio of theinput-shaft-side planetary gear set is low. In the case of trying tosecure a sufficient value of gear ratio of planetary gear set in thistechnique, another problem is caused that dimensions of the planetarygear set are upsized so as to upsize the outside dimensions of automatictransmission.

It is an object of the present invention to provide an automatictransmission that is capable of achieving seven forward speed-ratios bymeans of three simple planetary gear sets and five friction elements,and that is devised to reduce members passing on the radially outer sideof planetary gear set and/or devised to reduce the number of shaftspassing on the radially inner side of planetary gear set.

According to one aspect of the present invention, there is provided anautomatic transmission comprising: a first planetary gear set includinga first sun gear, a first pinion engaged with the first sun gear, afirst carrier supporting the first pinion, and a first ring gear engagedwith the first pinion and constantly locked; a second planetary gear setincluding a second sun gear, a second pinion engaged with the second sungear, a second carrier supporting the second pinion, and a second ringgear engaged with the second pinion; a third planetary gear setincluding a third sun gear, a third pinion engaged with the third sungear, a third carrier supporting the third pinion, the third carrierbeing connected with the second ring gear to define a rotating member,and a third ring gear engaged with the third pinion; an input shaftconstantly connected with the second carrier; an output shaft constantlyconnected with the rotating member; and five friction elements includinga first friction element adapted to selectively connect the firstcarrier with the third ring gear, a second friction element adapted toselectively connect the first sun gear with the second carrier, a thirdfriction element adapted to selectively connect the first sun gear withthe second sun gear, a fourth friction element adapted to selectivelyconnect the first carrier with the second carrier, and a fifth frictionelement adapted to selectively connect the first carrier with the secondsun gear, the automatic transmission being adapted to achieve at leastseven forward speed-ratios and one reverse speed-ratio, each of the atleast seven forward speed-ratios and one reverse speed-ratio beingachieved by an engaged state of two friction elements selected from thefive friction elements.

According to another aspect of the present invention, there is providedan automatic transmission comprising: a first planetary gear setincluding a first sun gear, a first pinion engaged with the first sungear, a first carrier supporting the first pinion, and a first ring gearengaged with the first pinion and constantly locked; a second planetarygear set including a second sun gear constantly locked, a second pinionengaged with the second sun gear, a second carrier supporting the secondpinion, and a second ring gear engaged with the second pinion, andconnected with the first carrier to define a rotating member; a thirdplanetary gear set including a third sun gear, a third pinion engagedwith the third sun gear, a third carrier supporting the third pinion,and a third ring gear engaged with the third pinion; an input shaftconstantly connected with the third carrier; an output shaft constantlyconnected with third ring gear; and five friction elements including afirst friction element adapted to selectively connect the second carrierwith the third ring gear, a second friction element adapted toselectively connect the first sun gear with the third carrier, a thirdfriction element adapted to selectively connect the first sun gear withthe third sun gear, a fourth friction element adapted to selectivelyconnect the rotating member with the third carrier, and a fifth frictionelement adapted to selectively connect the rotating member with thethird sun gear, the automatic transmission being adapted to achieve atleast seven forward speed-ratios and one reverse speed-ratio, each ofthe at least seven forward speed-ratios and one reverse speed-ratiobeing achieved by an engaged state of two friction elements selectedfrom the five friction elements.

According to still another aspect of the present invention, there isprovided an automatic transmission comprising: a first planetary gearset including a first sun gear, a first pinion engaged with the firstsun gear, a first carrier supporting the first pinion, and a first ringgear engaged with the first pinion; a second planetary gear setincluding a second sun gear, a second pinion engaged with the second sungear, a second carrier supporting the second pinion, and a second ringgear engaged with the second pinion and constantly locked; a thirdplanetary gear set including a third sun gear, a third pinion engagedwith the third sun gear, a third carrier supporting the third pinion,and a third ring gear engaged with the third pinion, the third ring gearbeing connected with the first carrier to define a first rotatingmember, the first ring gear being connected with the second carrier todefine a second rotating member; an input shaft constantly connectedwith the third carrier; an output shaft constantly connected with thefirst rotating member; and five friction elements including a firstfriction element adapted to stop a rotation of the first sun gear, asecond friction element adapted to selectively connect the third sungear with the third carrier, a third friction element adapted toselectively connect the second sun gear with the third sun gear, afourth friction element adapted to selectively connect the secondrotating member with the third carrier, and a fifth friction elementadapted to selectively connect the second rotating member with the thirdsun gear, the automatic transmission being adapted to achieve at leastseven forward speed-ratios and one reverse speed-ratio, each of the atleast seven forward speed-ratios and one reverse speed-ratio beingachieved by an engaged state of two friction elements selected from thefive friction elements.

According to still another aspect of the present invention, there isprovided an automatic transmission comprising: a first planetary gearset including a first sun gear, a first pinion engaged with the firstsun gear, a first carrier supporting the first pinion, and a first ringgear engaged with the first pinion; a second planetary gear setincluding a second sun gear, a second pinion engaged with the second sungear, a second carrier supporting the second pinion, and a second ringgear engaged with the second pinion; a third planetary gear setincluding a third sun gear constantly locked, a third pinion engagedwith the third sun gear, a third carrier supporting the third pinion,and a third ring gear engaged with the third pinion, the third ring gearbeing connected with the second carrier to define a first rotatingmember, the second ring gear being connected with the third carrier todefine a second rotating member; an input shaft constantly connectedwith the first carrier; an output shaft constantly connected with thefirst sun gear; and five friction elements including a first frictionelement adapted to selectively connect the first sun gear with thesecond rotating member, a second friction element adapted to selectivelyconnect the second sun gear with the first carrier, a third frictionelement adapted to selectively connect the second sun gear with thefirst ring gear, a fourth friction element adapted to selectivelyconnect the first rotating member with the first carrier, and a fifthfriction element adapted to selectively connect the second rotatingmember with the first ring gear, the automatic transmission beingadapted to achieve at least seven forward speed-ratios and one reversespeed-ratio, each of the at least seven forward speed-ratios and onereverse speed-ratio being achieved by an engaged state of two frictionelements selected from the five friction elements.

The other objects and features of this invention will become understoodfrom the following description with reference to the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a skeleton diagram showing an automatic transmission accordingto a first embodiment of the present invention.

FIG. 2 is a diagram showing a concrete example of an engagement table offriction elements and reduction gear ratios in the automatictransmission according to the first embodiment.

FIG. 3 is a view showing differences between respective first to seventhspeeds according to the first embodiment.

FIG. 4 is a skeleton diagram showing an automatic transmission accordingto a first modified example of the first embodiment.

FIG. 5 is a skeleton diagram showing an automatic transmission accordingto a second modified example of the first embodiment.

FIG. 6 is a skeleton diagram showing an automatic transmission accordingto a second embodiment of the present invention.

FIG. 7 is a diagram showing a concrete example of an engagement table offriction elements and reduction gear ratios in the automatictransmission according to the second embodiment.

FIG. 8 is a view showing differences between respective first to seventhspeeds according to the second embodiment.

FIG. 9 is a skeleton diagram showing an automatic transmission accordingto a first modified example of the second embodiment.

FIG. 10 is a skeleton diagram showing an automatic transmissionaccording to a third embodiment of the present invention.

FIG. 11 is a diagram showing a concrete example of an engagement tableof friction elements and reduction gear ratios in the automatictransmission according to the third embodiment.

FIG. 12 is a view showing differences between respective first toseventh speeds according to the third embodiment.

FIG. 13 is a skeleton diagram showing an automatic transmissionaccording to a fourth embodiment of the present invention.

FIG. 14 is a diagram showing a concrete example of an engagement tableof friction elements and reduction gear ratios in the automatictransmission according to the fourth embodiment.

DETAILED DESCRIPTION OF THE INVENTION

Reference will hereinafter be made to the drawings in order tofacilitate a better understanding of the present invention.

First Embodiment

At first, structures of a shift mechanism of a step automatictransmission according to a first embodiment of the present inventionwill now be explained. FIG. 1 is a skeleton diagram showing the shiftmechanism of the step (multiple-step type) automatic transmissionaccording to the first embodiment. FIG. 2 is a diagram showing aconcrete example of a table regarding engagements of friction elementsand reduction gear ratios (speed reducing ratios) in the automatictransmission according to the first embodiment.

The automatic transmission according to the first embodiment includes afirst planetary gear set PG1, a second planetary gear set PG2 and athird planetary gear set PG3, as a gear train, as shown in FIG. 1. Eachof the three planetary gear sets PG1, PG2 and PG3 is of single-piniontype. First planetary gear set PG1 includes a first sun gear S1, a firstring gear R1, and a first pinion P1 engaged or meshed with first sungear S1 and first ring gear R1. Second planetary gear set PG2 includes asecond sun gear S2, a second ring gear R2, and a second pinion P2engaged with second sun gear S2 and second ring gear R2. Third planetarygear set PG3 includes a third sun gear S3, a third ring gear R3, and athird pinion P3 engaged with third sun gear S3 and third ring gear R3.First, second and third pinions P1 to P3 are supported rotatablyrelative to a first carrier PC1, a second carrier PC2 and a thirdcarrier PC3, respectively. Namely, each carrier PC1, PC2 or PC3pivotally supports the corresponding pinion P1, P2 or P3.

An input shaft IN is always connected (or engaged) with second carrierPC2 (i.e., constantly rotates together with second carrier PC2). Secondring gear R2 is always connected with third carrier PC3 to form a firstrotating member M1. An output shaft OUT is always connected with firstrotating member M1. First ring gear R1 is always locked relative to atransmission case 1 (i.e., constantly fastened to the transmission case1). Third sun gear S3 is always locked relative to transmission case 1.

The automatic transmission further includes five clutches, i.e., firstto fifth friction elements A, B, C, D and E. The first friction elementA is provided between first carrier PC1 and third ring gear R3, and isadapted to selectively connect (engage) first carrier PC1 with thirdring gear R3. The second friction element B is provided between firstsun gear S1 and second carrier PC2, and is adapted to selectivelyconnect first sun gear S1 with second carrier PC2. The third frictionelement C is provided between first sun gear S1 and second sun gear S2,and is adapted to selectively connect first sun gear S1 with second sungear S2. The fourth friction element D is provided between first carrierPC1 and second carrier PC2, and is adapted to selectively connect firstcarrier PC1 with second carrier PC2. The fifth friction element E isprovided between first carrier PC1 and second sun gear S2, and isadapted to selectively connect first carrier PC1 with second sun gearS2.

Output shaft OUT is provided with an output gear or the like to transmitrotational driving force through a differential gear and a drive shaftto a drive wheel, which are not shown. In the case of the firstembodiment, since output shaft OUT is not obstructed by the other memberor the like, the automatic transmission is applicable to both of FFvehicle (front-engine front-drive vehicle) and FR vehicle (front-enginerear-drive vehicle).

The relations in engagements (connections) of the friction elementsunder respective speed-ratios (i.e., respective steps for shift) will beexplained below referring to the engagement table of FIG. 2 (theseengagements for respective speed-ratios are attained by a shift controlsection or means). In the table of FIG. 2, the sign ◯ represents theengagement (engaged state), and the blank represents the disengagement(released state).

At first, the states at the time of forward running will now beexplained. A first-speed (first speed-ratio) is achieved by engagingfirst friction element A and second friction element B. A second-speedis achieved by engaging first friction element A and third frictionelement C. A third-speed is achieved by engaging first friction elementA and fourth friction element D. A fourth-speed is achieved by engagingfirst friction element A and fifth friction element E. A fifth-speed isachieved by engaging fourth friction element D and fifth frictionelement E. A sixth-speed is achieved by engaging second friction elementB and fifth friction element E. A seventh-speed is achieved by engagingthird friction element C and fifth friction element E. Next, the stateat the time of reverse running is now explained. A reverse-speed isachieved by engaging third friction element C and fourth frictionelement D.

Next, a concrete example of the reduction gear ratios according to thefirst embodiment will now be explained referring to FIG. 2. Thefollowing explanations are given in the case where a gear ratioρ1=ZS1/ZR1 of first planetary gear set PG1 is equal to 0.45 (i.e.,ρ1=ZS1/ZR1=0.45), a gear ratio ρ2=ZS2/ZR2 of second planetary gear setPG2 is equal to 0.55 (i.e., ρ2=ZS2/ZR2=0.55), and a gear ratioρ3=ZS3/ZR3 of third planetary gear set PG3 is equal to 0.45 (i.e.,ρ3=ZS3/ZR3=0.45). Where, each of ZS1, ZS2, ZS3, ZR1, ZR2 and ZR3represents the number of teeth of the corresponding gear.

A reduction gear ratio i1 of the first-speed in the forward running isexpressed by a formula: i1=(1+ρ1)(1+ρ3)/ρ1. By assigning the concretenumerical values to this formula, reduction gear ratio i1 of the forwardfirst-speed is calculated as i1=4.672. The inverse of reduction gearratio i1 is equal to 0.214.

A reduction gear ratio i2 of the second-speed in the forward running isexpressed by a formula: i2=(ρ2(1+ρ1)(1+ρ3)+ρ1)/(ρ1(1+ρ2)). By assigningthe concrete numerical values to this formula, reduction gear ratio i2of the forward second-speed is calculated as i2=2.303. The inverse ofreduction gear ratio i2 is equal to 0.434.

A reduction gear ratio i3 of the third-speed in the forward running isexpressed by a formula: i3=1+ρ3. By assigning the concrete numericalvalues to this formula, reduction gear ratio i3 of the forwardthird-speed is calculated as i3=1.450. The inverse of reduction gearratio i3 is equal to 0.690.

A reduction gear ratio i4 of the fourth-speed in the forward running isexpressed by a formula: i4=(1+ρ2+ρ2 ρ3)/(1+ρ2). By assigning theconcrete numerical values to this formula, reduction gear ratio i4 ofthe forward fourth-speed is calculated as i4=1.160. The inverse ofreduction gear ratio i4 is equal to 0.862.

A reduction gear ratio i5 of the fifth-speed in the forward running isexpressed by a formula: i5=1.0. Without assigning the concrete numericalvalues to this formula, reduction gear ratio i5 of the forwardfifth-speed is equal to 1.000. The inverse of reduction gear ratio i5 isequal to 1.000.

A reduction gear ratio i6 of the sixth-speed in the forward running isexpressed by a formula: i6=(1+ρ1)/(1+ρ1+ρ2). By assigning the concretenumerical values to this formula, reduction gear ratio i6 of the forwardsixth-speed is calculated as i6=0.725. The inverse of reduction gearratio i6 is equal to 1.379.

A reduction gear ratio i7 of the seventh-speed in the forward running isexpressed by a formula: i7=1/(1+ρ2). By assigning the concrete numericalvalues to this formula, reduction gear ratio i7 of the forwardseventh-speed is calculated as i7=0.645. The inverse of reduction gearratio i7 is equal to 1.550.

A reduction gear ratio iR of the reverse-speed is expressed by aformula: iR=−ρ1/(ρ2−ρ1). By assigning the concrete numerical values tothis formula, reduction gear ratio iR of the reverse-speed is calculatedas iR=−4.500. The inverse of reduction gear ratio iR is equal to −0.222.

Next, a characteristic of the respective reduction gear ratios in thefirst embodiment is explained referring to FIG. 3. FIG. 3 is a tableshowing interrelations among the respective speed-ratios obtained by theautomatic transmission in the first embodiment. It is proper that aratio between speed-ratios (i.e., relation among respectivespeed-ratios, also called a step ratio) is evaluated based on how closea so-called V1000 is to arithmetic series (arithmetic progression). ThisV1000 is comparable to a vehicle-speed range which each speed-ratio isin charge of (a vehicle-speed range which each step for shift covers).That is, in the case where the V1000 is allocated in a manner ofarithmetic series, each of the in-charge vehicle-speed ranges inrespective speed-ratios has an equivalent width to one another. In thiscase, particularly; an upshift attains a rhythmical shift, and aselection of speed-ratio is conducted with no difficulty on a downhillor uphill road, so that the automatic transmission having a superiordrivability can be achieved.

In order to quantitatively determine how the V1000 has deviated ordeparted from the arithmetic series, the following procedure is used.That is, the inverses of reduction gear ratio values under therespective speed-ratios are normalized by regarding the inverse ofreduction gear ratio value under the highest-step speed-ratio(seventh-speed) as a value equal to 1, and then a deviation degree(amount) of these normalized values from the arithmetic series isquantitatively calculated. In the automatic transmission according tothe first embodiment, a standard deviation calculated is equal to 0.051which is recognized as an extremely small value.

Effects According to the First Embodiment

{circle around (1)} Effects by Virtue of Structural Skeleton as a Whole

In the first embodiment, the automatic transmission capable of attainingthe seven speeds of forward running and the one speed of reverse runningcan be realized with proper reduction gear ratios being ensured;although the automatic transmission is constructed by a limited numberof simple constructional elements, namely, the three sets of simple(single) planet gears (three single-pinion-type planetary gear sets) andthe five friction elements.

{circle around (2)} Effects by Virtue of the Usage of Three SimplePlanetary Gear Sets

Because of the usage of the three sets of simple planet gears (the usageof three single-pinion-type planetary gear sets), a gear noise and atransfer efficiency can be improved as compared to the case where doublepinions (double-pinion-type planetary gear set) are used. Further,because a diameter of the pinion does not need to be reduced in thisembodiment, a gear durability can be enhanced.

{circle around (3)} Effects Based on Gear Ratios

All of the gear ratios ρ1, ρ2 and ρ3 in the respective planetary gearsets are close to a center value 0.5. Accordingly, a possible rangewhich is obtained by freely setting the three gear ratios is wide sothat a degree of freedom in reduction gear ratio can become higher.

{circle around (4)} Effects Based on a Ratio Coverage in the ForwardRunning

A ratio coverage (gear-ratio width) of the forward running is defined bydividing the reduction gear ratio of the lowest-speed (step) by thereduction gear ratio of the highest-speed (step), i.e., the reductiongear ratio of the lowest-speed/the reduction gear ratio of thehighest-speed. It can be described that a compatibility between anaccelerating performance at the time of vehicle start and a fuel economyat the time of high speed cruise of vehicle becomes better, and also adegree of freedom to set the gear ratio value in respective forwardspeed-ratios becomes higher; as the value of ratio coverage becomesgreater. As concrete numerical values in the first embodiment, thereduction gear ratio of the forward first-speed is equal to 4.672 andthe reduction gear ratio of the forward seventh-speed is equal to 0.645.In this embodiment, the ratio coverage from first-speed to seventh-speedis equal to 7.24, and hence a sufficient ratio coverage can be ensured.Therefore, for example, the automatic transmission according to thefirst embodiment is useful also as a transmission for a vehicle equippedwith a diesel engine as its power source, although a width of rotationalspeed (number of revolutions) of diesel engine is narrower than that ofa gasoline engine and a torque of diesel engine is higher than that of agasoline engine having the same engine displacement.

Moreover, in the case where the gear ratio value of the low-speed sideis great relative to the ratio coverage, a torque transmitted to a finalgear becomes relatively great. Hence, this case requires a sufficientstrength of the automatic transmission or propeller shaft, so that thewhole of vehicle body is upsized. That is, it is preferable that thegear ratio value (value of speed ratio) of the lowest-speed is not sogreat under the same condition of ratio coverage. In an automatictransmission shown by FIG. 23 in the document of U.S. Pat. No.6,648,791, a gear ratio of the highest-speed (highest step for shift) isequal to 1. Hence in this technique, when trying to enlarge the ratiocoverage, a gear ratio (value) of the lowest-speed (lowest step forshift) needs to be enlarged, so that the upsizing of the automatictransmission and the propeller shaft is caused. On the other hand, inthe automatic transmission according to the first embodiment of thepresent invention, a sufficient ratio coverage can be ensured withoutthe necessity of enlarging the gear ratio of the lowest-speed so much.

{circle around (5)} Effects Based on 1-R Ratio

A value of 1-R ratio (Reverse-speed/First-speed) is a value near 1,concretely, equal to 0.963. Accordingly, an acceleration feel of vehiclerelative to a depressing adjustment of an accelerator pedal does notgreatly vary between at the time of forward running and at the time ofreverse running. Therefore, the problem that the drivability worsens canbe avoided.

{circle around (6)} Effects Based on Standard Deviation Related to theV1000

Since the standard deviation can be made an extremely small value (forexample, 0.051 in this embodiment), the widths of the vehicle-speedranges in the respective speed-ratios (the respective steps for shift)are equivalent to one another. Accordingly, particularly, the upshiftattains a rhythmical shift, and the selection of reduction gear ratio isconducted with no difficulty on a downhill or uphill road, so that theautomatic transmission having a superior drivability can be provided.

{circle around (7)}Effects Based on the Number of Changeovers Among theFriction Elements at the Time of Shift

(i) If one or more friction element is released and two or more frictionelements are engaged at the time of shift, or if two or more frictionelements are released and one or more friction element is engaged at thetime of shift; a torque control and a control for the engaging andreleasing timings of friction elements become complicated. Hence, from aviewpoint of avoidance of the complication of shift control, it isfavorable that one friction element is released and another frictionelement is engaged at the time of shift. That is, it is favorable that aso-called double-changeover is avoided. In the first embodiment, theshifts among the forward first-speed to the forward fourth-speed areperformed under the condition where first friction element A ismaintained in engaged state. Moreover, the shifts among the forwardfourth-speed to the forward seventh-speed are performed under thecondition where fifth friction element E is maintained in engaged state.Namely, each shift between adjacent two speed-ratios (gear steps) amongthe forward first-speed to seventh-speed can be achieved by releasingone friction element and by engaging one friction element. Accordingly,each of all the shifts between adjacent two speed-ratios of forwardrunning is performed by means of only the changeover from one frictionelement to the other one friction element. Therefore, the control duringthe shift can be prevented from being complicated.

(ii) As mentioned in the above (i), all the shifts between adjacent twospeed-ratios can be achieved by the changeover shift which releases onefriction element and engages one friction element. Moreover, similarly,each one-ratio-skip shift (all one-ratio-skip shifts, e.g., forwardfirst-speed→forward third-speed) among the forward first- toseventh-speeds can also be achieved by releasing one friction elementand by engaging the other one friction element. Accordingly, acontrollability thereof can be enhanced.

{circle around (8)}Effects Based on Layout

(i) In the automatic transmission according to the first embodiment, ona radially outer side of the three planetary gear sets, connectingmembers are disposed so as not to become in a three-layered form, asshown in the skeleton diagram of FIG. 1. That is, the number of theconnecting members radially covered or overlapped with each otherradially outside each planetary gear set is smaller than three.Accordingly, it becomes difficult to cause the retention of alubricating oil (disrupted flow of lubricating oil), so that the fueleconomy can be improved by reducing the frictions.

(ii) A member passing on a radially inner side of the three planetarygear sets is disposed in a single-shaft structure. That is, the numberof members (shafts) passing through a radially inside space (of the sungear) of each planetary gear set is equal to 1. Accordingly, dimensions(size) of each sun gear are not restricted as compared with thetechnique disclosed in the document of U.S. Pat. No. 6,648,791, and adegree of freedom to design the ratio between teeth numbers in eachplanetary gear set is high so that a degree of freedom to design theautomatic transmission can be enhanced.

(iii) Moreover, the rotating member passing on the outer peripheral sideof the planetary gear sets is formed in a single-layered structure, asshown by the skeleton diagram of FIG. 1. Generally in the automatictransmission, lubricating oil is always supplied to respective rotatingelements such as gears and bearings (not shown) for the purpose ofcooling, lubrication and the like. This lubricating oil is generallysupplied from a shaft-center side of the transmission by means ofcentrifugal force. At this time, if an efficiency of discharge(retrieving performance) of lubricating oil becomes worsened on theouter peripheral side of the planetary gear sets, oil temperature risesso that a durability of friction elements, bearings and the like isreduced. Since the rotating member passing on the outer peripheral sideof the planetary gear sets forms the single-layered structure in thefirst embodiment as mentioned above, the discharging efficiency oflubricating oil is not worsened so that the temperature rise issuppressed to improve the durability.

(iv) The automatic transmission according to the first embodiment can bedesigned to allow torque to be inputted to one side of the planetarygear sets and then to be outputted from another side of the planetarygear sets. Accordingly, the automatic transmission according to thefirst embodiment is applicable to both of a front-wheel drive vehicleand a rear-wheel drive vehicle, namely can be widely applied.

FIRST MODIFIED EXAMPLE

Next, a first modified example of the first embodiment according to thepresent invention will now be explained. Since a basic structure of thefirst modified example is the same as that of the pre-modified exampleexplained above, only structural parts of the first modified examplewhich are different from the above pre-modified example are nowexplained.

FIG. 4 is a skeleton diagram showing the first modified example. In theabove pre-modified example, all of first friction element A, thirdfriction element C, fourth friction element D and fifth friction elementE are located between first planetary gear set PG1 and second planetarygear set PG2. Contrary to this, in the first modified example, firstfriction element A is placed between second planetary gear set PG2 andthird planetary gear set PG3.

SECOND MODIFIED EXAMPLE

Next, a second modified example of the first embodiment according to thepresent invention will now be explained. Since a basic structure of thesecond modified example is the same as that of the above pre-modifiedexample of the first embodiment, only structural parts of the secondmodified example which are different from the pre-modified example arenow explained. In the above pre-modified example; the threesingle-pinion-type planetary gear sets, first planetary gear set PG1,second planetary gear set PG2 and third planetary gear set PG3 as thegear train are disposed in this order from the input side (firstplanetary gear set PG1→second planetary gear set PG2→third planetarygear set PG3). Contrary to this, in the second modified example, thethree planetary gear sets are disposed in order as second planetary gearset PG2→first planetary gear set PG1→third planetary gear set PG3 fromthe input side. FIG. 5 gives respective reference signs to the threeplanetary gear sets in the order as first planetary gear set PG1→secondplanetary gear set PG2→third planetary gear set PG3 from the left sideof FIG. 5. However, in a correspondence relation with the pre-modifiedexample; first planetary gear set PG1 of the second modified examplecorresponds to third planetary gear set PG3 of the pre-modified example,second planetary gear set PG2 of the second modified example correspondsto first planetary gear set PG1 of the pre-modified example, and thirdplanetary gear set PG3 of the second modified example corresponds tosecond planetary gear set PG2 of the pre-modified example.

In this modified example, the three planetary gear sets are arranged insuch a manner that second planetary gear set PG2 (which corresponds tofirst planetary gear set PG1 of the pre-modified example) is disposedaxially between third planetary gear set PG3 and first planetary gearset PG1 (which correspond to second planetary gear set PG2 and thirdplanetary gear set PG3 of the pre-modified example). Accordingly, afirst rotating member M1 composed of third ring gear R3 and firstcarrier PC1 (i.e., first rotating member M1 composed of second ring gearR2 and third carrier PC3 of the pre-modified example) covers the threeplanetary gear sets from the outer peripheral side thereof. Accordingly,the automatic transmission in the second modified example is mainlyapplied to a front-wheel drive vehicle.

All of the positional relations of respective friction elements, theconnecting relations of respective rotating members, and the engagementrelations of friction elements under the respective speed-ratios are thesame as those in the pre-modified example.

Second Embodiment

At first, structures according to a second embodiment of the presentinvention will now be explained. FIG. 6 is a skeleton diagram showingthe shift mechanism of a step (multiple-step type) automatictransmission according to the second embodiment. FIG. 7 is a diagramshowing a concrete example of a table regarding engagements of frictionelements and reduction gear ratios (speed reducing ratios) in theautomatic transmission according to the second embodiment.

The automatic transmission according to the second embodiment includes afirst planetary gear set PG1, a second planetary gear set PG2 and athird planetary gear set PG3, as a gear train, as shown in FIG. 6. Eachof the three planetary gear sets PG1, PG2 and PG3 is of single-piniontype. First planetary gear set PG1 includes a first sun gear S1, a firstring gear R1, and a first pinion P1 engaged or meshed with first sungear S1 and first ring gear R1. Second planetary gear set PG2 includes asecond sun gear S2, a second ring gear R2, and a second pinion P2engaged with second sun gear S2 and second ring gear R2. Third planetarygear set PG3 includes a third sun gear S3, a third ring gear R3, and athird pinion P3 engaged with third sun gear S3 and third ring gear R3.First, second and third pinions P1 to P3 are supported rotatablyrelative to a first carrier PC1, a second carrier PC2 and a thirdcarrier PC3, respectively. Namely, each carrier PC1, PC2 or PC3pivotally supports the corresponding pinion P1, P2 or P3.

An input shaft IN is always connected (or engaged) with third carrierPC3 (i.e., constantly rotates together with third carrier PC3). Secondring gear R2 is always connected with first carrier PC1 to form a firstrotating member M1. An output shaft OUT is always connected with thirdring gear R3. First ring gear R1 is always locked relative to atransmission case 1 (i.e., constantly fastened to the transmission case1). Second sun gear S2 is always locked relative to transmission case 1.

The automatic transmission further includes five clutches, i.e., firstto fifth friction elements A, B, C, D and E. The first friction elementA is provided between second carrier PC2 and third ring gear R3, and isadapted to selectively connect (engage) second carrier PC2 with thirdring gear RI. The second friction element B is provided between firstsun gear S1 and third carrier PC3, and is adapted to selectively connectfirst sun gear S1 with third carrier PC3. The third friction element Cis provided between first sun gear S1 and third sun gear S3, and isadapted to selectively connect first sun gear S1 with third sun gear S3.The fourth friction element D is provided between second ring gear R2and third carrier PC3, and is adapted to selectively connect second ringgear R2 with third carrier PC3. The fifth friction element E is providedbetween first carrier PC1 and third sun gear S3, and is adapted toselectively connect first carrier PC1 with third sun gear S3.

Output shaft OUT is provided with an output gear or the like to transmitrotational driving force through a differential gear and a drive shaftto a drive wheel, which are not shown. In the case of the secondembodiment, since output shaft OUT is obstructed by first ring gear R1,second sun gear S2 or the like, the automatic transmission is applicableto a FF vehicle (front-engine front-drive vehicle).

The relations in engagements (connections) of the friction elementsunder respective speed-ratios (i.e., respective steps for shift) will beexplained below referring to the engagement table of FIG. 7 (theseengagements for respective speed-ratios are attained by a shift controlsection or means). In the table of FIG. 7, the sign ◯ represents theengagement, and the blank represents the disengagement.

At first, the states at the time of forward running will now beexplained. A first-speed (first speed-ratio) is achieved by engagingfirst friction element A and second friction element B. A second-speedis achieved by engaging first friction element A and third frictionelement C. A third-speed is achieved by engaging first friction elementA and fourth friction element D. A fourth-speed is achieved by engagingfirst friction element A and fifth friction element E. A fifth-speed isachieved by engaging fourth friction element D and fifth frictionelement E. A sixth-speed is achieved by engaging second friction elementB and fifth friction element E. A seventh-speed is achieved by engagingthird friction element C and fifth friction element E. Next, the stateat the time of reverse running is now explained. A reverse-speed isachieved by engaging third friction element C and fourth frictionelement D.

Next, a concrete example of the reduction gear ratios according to thesecond embodiment will now be explained referring to FIG. 7. Thefollowing explanations are given in the case where a gear ratioρ1=ZS1/ZR1 of first planetary gear set PG1 is equal to 0.55 (i.e.,ρ1=ZS1/ZR1=0.55), a gear ratio ρ2=ZS2/ZR2 of second planetary gear setPG2 is equal to 0.50 (i.e., ρ2=ZS2/ZR2=0.50), and a gear ratioρ3=ZS3/ZR3 of third planetary gear set PG3 is equal to 0.65 (i.e.,ρ3=ZS3/ZR3=0.65). Where, each of ZS1, ZS2, ZS3, ZR1, ZR2 and ZR3represents the number of teeth of the corresponding gear.

A reduction gear ratio i1 of the first-speed in the forward running isexpressed by a formula: i1=(1+ρ1)(1+ρ2)/ρ1. By assigning the concretenumerical values to this formula, reduction gear ratio i1 of the forwardfirst-speed is calculated as i1=4.227. The inverse of reduction gearratio i1 is equal to 0.237.

A reduction gear ratio i2 of the second-speed in the forward running isexpressed by a formula: i2=(ρ1+ρ3(1+ρ1)(1+ρ2))/(ρ1(1+ρ3)). By assigningthe concrete numerical values to this formula, reduction gear ratio i2of the forward second-speed is calculated as i2=2.271. The inverse ofreduction gear ratio i2 is equal to 0.440.

A reduction gear ratio i3 of the third-speed in the forward running isexpressed by a formula: i3=1+ρ2. By assigning the concrete numericalvalues to this formula, reduction gear ratio i3 of the forwardthird-speed is calculated as i3=1.500. The inverse of reduction gearratio i3 is equal to 0.667.

A reduction gear ratio i4 of the fourth-speed in the forward running isexpressed by a formula: i4=(1+ρ3+ρ2 ρ3)/(1+ρ3). By assigning theconcrete numerical values to this formula, reduction gear ratio i4 ofthe forward fourth-speed is calculated as i4=1.197. The inverse ofreduction gear ratio i4 is equal to 0.835.

A reduction gear ratio i5 of the fifth-speed in the forward running isexpressed by a formula: i5=1.0. Without assigning the concrete numericalvalues to this formula, reduction gear ratio i5 of the forwardfifth-speed is equal to 1.000. The inverse of reduction gear ratio i5 isequal to 1.000.

A reduction gear ratio i6 of the sixth-speed in the forward running isexpressed by a formula: i6=(1+ρ1)/(1+ρ1+ρ3). By assigning the concretenumerical values to this formula, reduction gear ratio i6 of the forwardsixth-speed is calculated as i6=0.705. The inverse of reduction gearratio i6 is equal to 1.418.

A reduction gear ratio i7 of the seventh-speed in the forward running isexpressed by a formula: i7=1/(1+ρ3). By assigning the concrete numericalvalues to this formula, reduction gear ratio i7 of the forwardseventh-speed is calculated as i7=0.606. The inverse of reduction gearratio i7 is equal to 1.650.

A reduction gear ratio iR of the reverse-speed is expressed by aformula: iR=−ρ1/(ρ3−ρ1). By assigning the concrete numerical values tothis formula, reduction gear ratio iR of the reverse-speed is calculatedas iR=−5.500. The inverse of reduction gear ratio iR is equal to −0.182.

Next, a characteristic of the respective reduction gear ratios in thesecond embodiment is explained referring to FIG. 8. FIG. 8 is a tableshowing interrelations among the respective speed-ratios obtained by theautomatic transmission in the second embodiment. It is proper that aratio between speed-ratios (i.e., relations among respectivespeed-ratios) is evaluated based on how close the V1000 is to arithmeticseries (arithmetic progression). This V1000 is comparable to avehicle-speed range which each speed-ratio is in charge of (avehicle-speed range which each speed-ratio covers). That is, in the casewhere the V1000 is allocated in a manner of arithmetic series, each ofthe in-charge vehicle-speed ranges in respective speed-ratios has anequivalent width to one another. In this case, particularly; upshiftscan become rhythmical, and a selection of speed-ratio is conducted withno difficulty on a downhill or uphill road, so that the automatictransmission having a superior drivability can be achieved.

In order to quantitatively determine how the V1000 has deviated ordeparted from the arithmetic series, the following procedure is used.That is, the inverses of reduction gear ratio values under therespective speed-ratios are normalized by regarding the inverse ofreduction gear ratio value under the highest-step speed-ratio(seventh-speed) as a value equal to 1, and then a deviation degree(amount) of these normalized values from the arithmetic series isquantitatively calculated. In the automatic transmission according tothe second embodiment, the standard deviation calculated is equal to0.052 which is recognized as an extremely small value.

Effects According to the Second Embodiment

{circle around (1)} CD Effects by Virtue of Structural Skeleton as aWhole

In the second embodiment, the automatic transmission capable ofattaining the seven speeds of forward running and the one speed ofreverse running can be realized with proper reduction gear ratios beingensured; although the automatic transmission is constructed by a limitednumber of simple constructional elements, namely, the three sets ofsimple (single) planet gears (three single-pinion-type planetary gearsets) and the five friction elements.

{circle around (2)} Effects by Virtue of the Usage of Three SimplePlanetary Gear Sets

Because of the usage of the three sets of simple planet gears (the usageof three single-pinion-type planetary gear sets), a gear noise and atransfer efficiency can be improved as compared to the case where doublepinions (double-pinion-type planetary gear set) are used. Further,because a diameter of the pinion does not need to be reduced in thisembodiment, the gear durability can be enhanced.

{circle around (3)} Effects Based on Gear Ratio

All of the gear ratios ρ1, ρ2 and ρ3 in the respective planetary gearsets are close to a center value 0.5. Accordingly, a possible rangewhich is obtained by freely setting the three gear ratios is wide sothat a degree of freedom in reduction gear ratio can become higher.

{circle around (4)} Effects Based on Ratio Coverage in the ForwardRunning

A ratio coverage (gear-ratio width) of the forward running is defined bydividing the reduction gear ratio of the lowest-speed (step) by thereduction gear ratio of the highest-speed (step), i.e., the reductiongear ratio of the lowest-speed/the reduction gear ratio of thehighest-speed. It can be described that a compatibility between anaccelerating performance at the time of vehicle start and a fuel economyat the time of high speed cruise of vehicle becomes better, and also adegree of freedom to set the gear ratio in respective forwardspeed-ratios becomes higher; as the value of ratio coverage becomesgreater. As concrete numerical values in the second embodiment, thereduction gear ratio of the forward first-speed is equal to 4.227 andthe reduction gear ratio of the forward seventh-speed is equal to 0.606.In this embodiment, the ratio coverage from first-speed to seventh-speedis equal to 6.98, and hence a sufficient ratio coverage can be ensured.Therefore, for example, the automatic transmission according to thesecond embodiment is useful also as a transmission for a vehicleequipped with a diesel engine as its power source, although a width ofrotational speed (number of revolutions) of diesel engine is narrowerthan that of a gasoline engine and a torque of diesel engine is higherthan that of a gasoline engine having the same engine displacement.

Moreover, in the case where the gear ratio of the low-speed side isgreat relative to the ratio coverage, a torque transmitted to the finalgear becomes relatively great. Hence, this case requires a sufficientstrength of the automatic transmission or propeller shaft, so that thewhole of vehicle body is upsized. That is, it is preferable that thegear ratio (value of speed ratio) of the lowest-speed is not so greatunder the same condition of ratio coverage. In an automatic transmissionshown by FIG. 23 in the document of U.S. Pat. No. 6,648,791, a gearratio of the highest-speed (highest step for shift) is equal to 1. Hencein this technique, when trying to enlarge the ratio coverage, a gearratio (value) of the lowest-speed (lowest step for shift) needs to beenlarged, so that the upsizing of the automatic transmission and thepropeller shaft is caused. On the other hand, in the automatictransmission according to the second embodiment of the presentinvention, a sufficient ratio coverage can be ensured without thenecessity of enlarging the gear ratio of the lowest-speed so much.

{circle around (5)} Effects Based on 1-R Ratio

A value of 1-R ratio (Reverse-speed/First-speed) is a value near 1,concretely, equal to 1.30. Accordingly, an acceleration feel of vehiclerelative to a depressing adjustment of an accelerator pedal does notgreatly vary between at the time of forward running and at the time ofreverse running. Therefore, the problem that the drivability is worsenedcan be avoided.

{circle around (6)} Effects Based on Standard Deviation Related to theV1000

Since the standard deviation can be made an extremely small value (forexample, 0.052 in this embodiment), the widths of the vehicle-speedranges in the respective speed-ratios (the respective steps for shift)are equivalent to one another. Accordingly, particularly, the upshiftattains a rhythmical shift, and the selection of reduction gear ratio isconducted with no difficulty on a downhill or uphill road, so that theautomatic transmission having a superior drivability can be provided.

{circle around (7)} Effects Based on the Number of Changeovers Among theFriction Elements at the Time of Shift

(i) If one or more friction element is released and two or more frictionelements are engaged at the time of shift, or if two or more frictionelements are released and one or more friction element is engaged at thetime of shift; a torque control and a control for the engaging andreleasing timings of friction elements become complicated. Hence, from aviewpoint of avoidance of the complication of shift control, it isfavorable that one friction element is released and the other onefriction element is engaged at the time of shift. That is, it isfavorable that a so-called double-changeover is avoided. In the secondembodiment, the shifts among the forward first-speed to the forwardfourth-speed are performed under the condition where first frictionelement A is maintained in engaged state. Moreover, the shifts among theforward fourth-speed to the forward seventh-speed are performed underthe condition where fifth friction element E is maintained in engagedstate. Namely, each shift between adjacent two speed-ratios (gear steps)among the forward first- to seventh-speeds can be achieved by releasingone friction element and by engaging one friction element. Accordingly,each of all the shifts between adjacent two speed-ratios of forwardrunning is performed by means of only the changeover from one frictionelement to the other one friction element. Therefore, the control duringthe shift can be prevented from being complicated.

(ii) As mentioned in the above (i), all the shifts between adjacent twospeed-ratios can be achieved by the changeover shift which releases onefriction element and engages one friction element. Moreover, similarly,each one-ratio-skip shift (e.g., forward first-speed→forwardthird-speed) among the forward first- to seventh-speeds can also beachieved by releasing one friction element and by engaging the other onefriction element. Accordingly, a controllability thereof can beenhanced.

{circle around (8)} Effects Based on Layout

(i) In the automatic transmission according to the second embodiment, ona radially outer side of the three planetary gear sets, connectingmembers are disposed so as not to become in a three-layered form, asshown in the skeleton diagram of FIG. 6. Accordingly, the automatictransmission becomes resistant to the occurrence of the retention oflubricating oil (disrupted flow of lubricating oil), so that the fueleconomy can be improved by reducing the frictions.

(ii) Moreover, the rotating member passing on the outer peripheral sideof the planetary gear sets is formed in a single-layered structure, asshown by the skeleton diagram of FIG. 6. Generally in the automatictransmission, lubricating oil is always supplied to respective rotatingelements such as gears and bearings (not shown) for the purpose ofcooling, lubrication and the like. This lubricating oil is generallysupplied from a shaft-center side of the transmission by means ofcentrifugal force. At this time, if an efficiency of discharge oflubricating oil becomes worsened on the outer peripheral side of theplanetary gear sets, oil temperature rises so that a durability offriction elements, bearings and the like is reduced. Since the rotatingmember passing on the outer peripheral side of the planetary gear setsforms the single-layered structure in the second embodiment as mentionedabove, the discharging efficiency of lubricating oil is not worsened sothat the temperature rise is suppressed to improve the durability.

FIRST MODIFIED EXAMPLE

Next, a first modified example of the second embodiment according to thepresent invention will now be explained. Since a basic structure of thefirst modified example is the same as that of the pre-modified exampleof the second embodiment explained above, only structural parts of thefirst modified example which are different from the above pre-modifiedexample are now explained. In the above pre-modified example of thesecond embodiment; the three single-pinion-type planetary gear sets,third planetary gear set PG3, second planetary gear set PG2 and firstplanetary gear set PG1 as the gear train are disposed in this order(third planetary gear set PG3→second planetary gear set PG2→firstplanetary gear set PG1) from a side of planetary gear set connected withinput shaft IN and output shaft OUT. Contrary to this, in the firstmodified example, the three planetary gear sets are disposed in order asthird planetary gear set PG3→first planetary gear set PG1→secondplanetary gear set PG2 from the side of planetary gear set connectedwith input shaft IN and output shaft OUT. FIG. 9 gives respectivereference signs to the three planetary gear sets in the order as firstplanetary gear set PG1→second planetary gear set PG2→third planetarygear set PG3 from the left side of FIG. 9. However, in a correspondencerelation with the pre-modified example of the second embodiment; firstplanetary gear set PG1 of the first modified example corresponds tosecond planetary gear set PG2 of the pre-modified example, secondplanetary gear set PG2 of the first modified example corresponds tofirst planetary gear set PG1 of the pre-modified example, and thirdplanetary gear set PG3 of the first modified example corresponds tothird planetary gear set PG3 of the pre-modified example.

All of the positional relations of respective friction elements, theconnecting relations of respective rotating members, and the engagementrelations of friction elements under the respective speed-ratios are thesame as those in the pre-modified example.

In this modified example, the three planetary gear sets are arranged insuch a manner that first planetary gear set PG1 (which corresponds tosecond planetary gear set PG2 of the pre-modified example of the secondembodiment) is disposed at an end portion of the three planetary gearsets, and thereby first ring gear R1 (which corresponds to second ringgear R2 of the pre-modified example) can be caught from its radiallyinner side by a first rotating member M1 connecting first ring gear R1with second carrier PC2 (which corresponds first rotating member M1connecting second ring gear R2 with first carrier PC1 of thepre-modified example). Accordingly, in this modified example, firstrotating member M1 can be disposed on the outer peripheral side of theplanetary gear sets, and thereby the rotating shaft passing throughradially inner side of the sun gears can be designed in a double-shaftstructure at a maximum. Therefore, dimensions of each sun gear are notrestricted as compared with the technique disclosed in the document ofU.S. Pat. No. 6,648,791, and a degree of freedom to design the ratiobetween teeth numbers in each planetary gear set is high so that adegree of freedom to design the automatic transmission can be enhanced.

Third Embodiment

At first, structures according to a third embodiment of the presentinvention will now be explained. FIG. 10 is a skeleton diagram showingthe shift mechanism of a step (multiple-step type) automatictransmission according to the third embodiment. FIG. 11 is a diagramshowing a concrete example of a table regarding engagements of frictionelements and reduction gear ratios (speed reducing ratios) in theautomatic transmission according to the third embodiment.

The automatic transmission according to the third embodiment includes afirst planetary gear set PG1, a second planetary gear set PG2 and athird planetary gear set PG3, as a gear train, as shown in FIG. 10. Eachof the three planetary gear sets PG1, PG2 and PG3 is of single-piniontype. First planetary gear set PG1 includes a first sun gear 51, a firstring gear R1, and a first pinion P1 engaged or meshed with first sungear S1 and first ring gear R1. Second planetary gear set PG2 includes asecond sun gear S2, a second ring gear R2, and a second pinion P2engaged with second sun gear S2 and second ring gear R2. Third planetarygear set PG3 includes a third sun gear S3, a third ring gear R3, and athird pinion P3 engaged with third sun gear S3 and third ring gear R3.First, second and third pinions P1 to P3 are supported rotatablyrelative to a first carrier PC1, a second carrier PC2 and a thirdcarrier PC3, respectively. Namely, each carrier PC1, PC2 or PC3pivotally supports the corresponding pinion P1, P2 or P3.

An input shaft IN is always connected with third carrier PC3 (i.e.,constantly rotates together with third carrier PC3). Third ring gear R3is always connected with first carrier PC1 by a first rotating memberM1. An output shaft OUT is always connected with first rotating memberM1, namely forms first rotating member M1. First ring gear R1 is alwaysconnected with second carrier PC2 to form a second rotating member M2.Second ring gear R2 is always locked relative to a transmission case 1(i.e., constantly fastened to the transmission case 1).

The automatic transmission further includes one brake, i.e., a firstfriction element A; and four clutches, i.e., second to fifth frictionelements B, C, D and E. The first friction element A is provided betweenfirst sun gear 51 and transmission case 1, and is adapted to selectivelylock (stop) a rotation of first sun gear 51 relative to transmissioncase 1 (i.e., selectively fasten first sun gear S1 to transmission case1). The second friction element B is provided between third sun gear S3and third carrier PC3, and is adapted to selectively connect third sungear S3 with third carrier PC3. The third friction element C is providedbetween second sun gear S2 and third sun gear S3, and is adapted toselectively connect second sun gear S2 with third sun gear S3. Thefourth friction element D is provided between second rotating member M2and third carrier PC3, and is adapted to selectively connect secondrotating member M2 with third carrier PC3. The fifth friction element Eis provided between second rotating member M2 and third sun gear S3, andis adapted to selectively connect second rotating member M2 with thirdsun gear S3. In the third embodiment, the three planetary gear sets aredisposed in order as third planetary gear set PG3=second planetary gearset PG2→first planetary gear set PG1 from the side of planetary gear setconnected with input shaft IN and output shaft OUT.

Output shaft OUT is provided with an output gear or the like to transmitrotational driving force through a differential gear and a drive shaftto a drive wheel, which are not shown. In the case of the thirdembodiment, since output shaft OUT is obstructed by a member for lockingsecond ring gear R2, the automatic transmission is applicable to a FFvehicle.

The relations in engagements (connections) of the friction elementsunder respective speed-ratios (i.e., respective steps for shift) will beexplained below referring to the engagement table of FIG. 11 (theseengagements for respective speed-ratios are attained by a shift controlsection or means). In the table of FIG. 11, the sign ◯ represents theengagement, and the blank represents the disengagement.

At first, the states at the time of forward running will now beexplained. A first-speed (first speed-ratio) is achieved by engagingfirst friction element A and second friction element B. A second-speedis achieved by engaging first friction element A and third frictionelement C. A third-speed is achieved by engaging first friction elementA and fourth friction element D. A fourth-speed is achieved by engagingfirst friction element A and fifth friction element E. A fifth-speed isachieved by engaging fourth friction element D and fifth frictionelement E. A sixth-speed is achieved by engaging second friction elementB and fifth friction element E. A seventh-speed is achieved by engagingthird friction element C and fifth friction element E. Next, the stateat the time of reverse running is now explained. A reverse-speed isachieved by engaging third friction element C and fourth frictionelement D.

Next, a concrete example of the reduction gear ratios according to thethird embodiment will now be explained referring to FIG. 11. Thefollowing explanations are given in the case where a gear ratioρ1=ZS1/ZR1 of first planetary gear set PG1 is equal to 0.45 (i.e.,ρ1=ZS1/ZR1=0.45), a gear ratio ρ2=ZS2/ZR2 of second planetary gear setPG2 is equal to 0.45 (i.e., ρ2=ZS2/ZR2=0.45), and a gear ratioρ3=ZS3/ZR3 of third planetary gear set PG3 is equal to 0.55 (i.e.,ρ3=ZS3/ZR3=0.55). Where, each of ZS1, ZS2, ZS3, ZR1, ZR2 and ZR3represents the number of teeth of the corresponding gear.

A reduction gear ratio i1 of the first-speed in the forward running isexpressed by a formula: i1=(1+ρ1)(1+ρ2)/ρ1. By assigning the concretenumerical values to this formula, reduction gear ratio i1 of the forwardfirst-speed is calculated as i1=4.672. The inverse of reduction gearratio i1 is equal to 0.214.

A reduction gear ratio i2 of the second-speed in the forward running isexpressed by a formula: i2=(ρ2+ρ3(1+ρ1)(1+ρ2))/(ρ2(1+ρ3)). By assigningthe concrete numerical values to this formula, reduction gear ratio i2of the forward second-speed is calculated as i2=2.303. The inverse ofreduction gear ratio i2 is equal to 0.434.

A reduction gear ratio i3 of the third-speed in the forward running isexpressed by a formula: i3=1+ρ1. By assigning the concrete numericalvalues to this formula, reduction gear ratio i3 of the forwardthird-speed is calculated as i3=1.450. The inverse of reduction gearratio i3 is equal to 0.690.

A reduction gear ratio i4 of the fourth-speed in the forward running isexpressed by a formula: i4=(1+ρ3+ρ1 ρ3)/(1+ρ3). By assigning theconcrete numerical values to this formula, reduction gear ratio i4 ofthe forward fourth-speed is calculated as i4=1.160. The inverse ofreduction gear ratio i4 is equal to 0.862.

A reduction gear ratio i5 of the fifth-speed in the forward running isexpressed by a formula: i5=1.0. Without assigning the concrete numericalvalues to this formula, reduction gear ratio i5 of the forwardfifth-speed is equal to 1.000. The inverse of reduction gear ratio i5 isequal to 1.000.

A reduction gear ratio i6 of the sixth-speed in the forward running isexpressed by a formula: i6=(1+ρ2)/(1+ρ2+ρ3). By assigning the concretenumerical values to this formula, reduction gear ratio i6 of the forwardsixth-speed is calculated as i6=0.725. The inverse of reduction gearratio i6 is equal to 1.379.

A reduction gear ratio i7 of the seventh-speed in the forward running isexpressed by a formula: i7=1/(1+ρ3). By assigning the concrete numericalvalues to this formula, reduction gear ratio i7 of the forwardseventh-speed is calculated as i7=0.645. The inverse of reduction gearratio i7 is equal to 1.550.

A reduction gear ratio iR of the reverse-speed is expressed by aformula: iR=ρ2/(ρ2−ρ3). By assigning the concrete numerical values tothis formula, reduction gear ratio iR of the reverse-speed is calculatedas iR=−4.500. The inverse of reduction gear ratio iR is equal to −0.222.

Next, a characteristic of the respective reduction gear ratios in thethird embodiment is explained referring to FIG. 12. FIG. 12 is a tableshowing interrelations among the respective speed-ratios obtained by theautomatic transmission in the third embodiment. It is proper that aratio between speed-ratios (i.e., relations among respectivespeed-ratios) is evaluated based on how close the V1000 is to arithmeticseries. This V1000 is comparable to a vehicle-speed range which eachspeed-ratio is in charge of (i.e., a vehicle-speed range which each stepfor shift covers). That is, in the case where the V1000 is allocated ina manner of arithmetic series, each of the in-charge vehicle-speedranges in respective speed-ratios has an equivalent width to oneanother. In this case, particularly; upshifts can become rhythmical, anda selection of speed-ratio is conducted with no difficulty on a downhillor uphill road, so that the automatic transmission having a superiordrivability can be achieved.

In order to quantitatively determine how the V1000 has deviated ordeparted from the arithmetic series, the following procedure is used.That is, the inverses of reduction gear ratio values under therespective speed-ratios are normalized by regarding the inverse ofreduction gear ratio value under the highest-step speed-ratio(seventh-speed) as a value equal to 1, and then a deviation degree(amount) of these normalized values from the arithmetic series isquantitatively calculated. In the automatic transmission according tothe third embodiment, the standard deviation calculated is equal to0.051 which is recognized as an extremely small value.

Effects according to the Third Embodiment

{circle around (1)} Effects by Virtue of Structural Skeleton as a Whole

In the third embodiment, the automatic transmission capable of attainingthe seven speeds of forward running and the one speed of reverse runningcan be realized with proper reduction gear ratios being ensured;although the automatic transmission is constructed by a limited numberof simple constructional elements, namely, the three sets of simple(single) planet gears (three single-pinion-type planetary gear sets) andthe five friction elements.

{circle around (2)} Effects by Virtue of the Usage of Three SimplePlanetary Gear Sets

Because of the usage of the three sets of simple planet gears (the usageof three single-pinion-type planetary gear sets), a gear noise and atransfer efficiency can be improved as compared to the case where doublepinions (double-pinion-type planetary gear set) are used. Further,because a diameter of the pinion does not need to be reduced in thisembodiment, the gear durability can be enhanced.

{circle around (3)} Effects Based on Gear Ratio

All of the gear ratios ρ1, ρ2 and ρ3 in the respective planetary gearsets are close to a center value 0.5. Accordingly, a possible rangewhich is obtained by freely setting the three gear ratios is wide sothat a degree of freedom in reduction gear ratio can become higher.

{circle around (4)} Effects Based on a Ratio Coverage in the ForwardRunning

A ratio coverage (gear-ratio width) of the forward running is defined bydividing the reduction gear ratio of the lowest-speed (step) by thereduction gear ratio of the highest-speed, i.e., the reduction gearratio of the lowest-speed/the reduction gear ratio of the highest-speed.A compatibility between an accelerating performance at the time ofvehicle start and a fuel economy at the time of high speed cruise ofvehicle becomes better, and also a degree of freedom to set the gearratio value in each forward speed-ratio becomes higher; as the value ofratio coverage becomes greater. As concrete numerical values in thethird embodiment, the reduction gear ratio of the forward first-speed isequal to 4.672 and the reduction gear ratio of the forward seventh-speedis equal to 0.645. In this embodiment, the ratio coverage fromfirst-speed to seventh-speed is equal to 7.24, and hence a sufficientratio coverage can be ensured. Therefore, for example, the automatictransmission according to the third embodiment is useful also as atransmission for a vehicle equipped with a diesel engine as its powersource, although a width of rotational speed (number of revolutions) ofdiesel engine is narrower than that of a gasoline engine and a torque ofdiesel engine is higher than that of a gasoline engine having the sameengine displacement.

Moreover, in the case where the gear ratio value of the low-speed sideis great relative to the ratio coverage, a torque transmitted to a finalgear becomes relatively great. Hence, this case requires a sufficientstrength of the automatic transmission or propeller shaft, so that thewhole of vehicle body is upsized. That is, it is preferable that thegear ratio value (value of speed ratio) of the lowest-speed is not sogreat under the same condition of ratio coverage. In an automatictransmission shown by FIG. 23 in the document of U.S. Pat. No.6,648,791, a gear ratio of the highest-speed (highest step for shift) isequal to 1. Hence in this technique, when trying to enlarge the ratiocoverage, a gear ratio (value) of the lowest-speed (lowest step forshift) needs to be enlarged, so that the upsizing of the automatictransmission and the propeller shaft is caused. On the other hand, inthe automatic transmission according to the third embodiment of thepresent invention, a sufficient ratio coverage can be ensured withoutthe necessity of enlarging the gear ratio of the lowest-speed so much.

{circle around (5)} Effects Based on 1-R Ratio

A value of 1-R ratio (Reverse-speed/First-speed) is a value near 1,concretely, equal to 0.963. Accordingly, an acceleration feel of vehiclerelative to a depressing adjustment of an accelerator pedal does notgreatly vary between at the time of forward running and at the time ofreverse running. Therefore, the problem that the drivability is worsenedcan be avoided.

{circle around (6)} Effects Based on Standard Deviation Related to theV1000

Since the standard deviation can be made an extremely small value(concretely, 0.051 in this embodiment), the widths of the vehicle-speedranges in the respective speed-ratios (the respective steps for shift)are equivalent to one another. Accordingly, particularly, the upshiftsis rhythmical, and the selection of reduction gear ratio is carried outwith no difficulty on a downhill or uphill road, so that the automatictransmission having a superior drivability can be provided.

{circle around (7)} Effects Based on the Number of Changeovers Among theFriction Elements at the Time of Shift

(i) If one or more friction element is released and two or more frictionelements are engaged at the time of shift, or if two or more frictionelements are released and one or more friction element is engaged at thetime of shift; a torque control and a control for the engaging andreleasing timings of friction elements become complicated. Hence, from aviewpoint of avoidance of the complication of shift control, it isfavorable that one friction element is released and another frictionelement is engaged at the time of shift. That is, it is favorable that aso-called double-changeover is avoided. In the third embodiment, theshifts among the forward first-speed to the forward fourth-speed areperformed under the condition where first friction element A ismaintained in engaged state. Moreover, the shifts among the forwardfourth-speed to the forward seventh-speed are performed under thecondition where fifth friction element E is maintained in engaged state.Namely, each shift between adjacent two speed-ratios among the forwardfirst- to seventh-speeds can be achieved by releasing one frictionelement and by engaging one friction element. Accordingly, each of allthe shifts between adjacent two speed-ratios of forward running isperformed by means of only the changeover from one friction element tothe other one friction element. Therefore, the control during the shiftcan be prevented from being complicated.

(ii) As mentioned in the above (i), all the shifts between adjacent twospeed-ratios can be achieved by the changeover shift which releases onefriction element and engages one friction element. Moreover, similarly,each one-ratio-skip shift (e.g., forward first-speed→forwardthird-speed) among the forward first- to seventh-speeds can also beachieved by releasing one friction element and by engaging the other onefriction element. Accordingly, a controllability thereof can beenhanced.

{circle around (8)} Effects Based on Layout

(i) In the automatic transmission according to the third embodiment, ona radially outer side of the three planetary gear sets, connectingmembers are disposed so as not to become in a three-layered form, asshown in the skeleton diagram of FIG. 10. Accordingly, it is suppressedthat the retention of a lubricating oil (disrupted flow of lubricatingoil) is caused, so that the fuel economy can be improved by reducing thefriction.

(ii) A member passing on a radially inner side of the three planetarygear sets is disposed in a single-shaft structure. Accordingly,dimensions of each sun gear are not restricted as compared with thetechnique disclosed in the document of U.S. Pat. No. 6,648,791, and adegree of freedom to design the ratio between teeth numbers in eachplanetary gear set is high so that a degree of freedom to design theautomatic transmission can be enhanced.

{circle around (9)} Effects from a Viewpoint of the Number of FrictionElements

In the third embodiment, first friction element A is provided as abrake. That is, since the five friction elements include the brake, theincrease of the number of sealings for rotation and the increase ofcentrifugal canceling mechanisms can be suppressed as compared with thecase where the number of clutches is large. Thereby, the increase of thenumber of components and the increase of axial length can be suppressedwhile enhancing the fuel economy.

Fourth Embodiment

At first, structures according to a fourth embodiment of the presentinvention will now be explained. FIG. 13 is a skeleton diagram showingthe shift mechanism of a step (multiple-step type) automatictransmission according to the fourth embodiment. FIG. 13 is a diagramshowing a concrete example of a table regarding engagements of frictionelements and reduction gear ratios (speed reducing ratios) in theautomatic transmission according to the fourth embodiment.

The automatic transmission according to the fourth embodiment includes afirst planetary gear set PG1, a second planetary gear set PG2 and athird planetary gear set PG3, as a gear train, as shown in FIG. 13. Eachof the three planetary gear sets PG1, PG2 and PG3 is of single-piniontype. First planetary gear set PG1 includes a first sun gear S1, a firstring gear R1, and a first pinion P1 engaged or meshed with first sungear 51 and first ring gear R1. Second planetary gear set PG2 includes asecond sun gear S2, a second ring gear R2, and a second pinion P2engaged with second sun gear S2 and second ring gear R2. Third planetarygear set PG3 includes a third sun gear S3, a third ring gear R3, and athird pinion P3 engaged with third sun gear S3 and third ring gear R3.First, second and third pinions P1 to P3 are supported rotatablyrelative to a first carrier PC1, a second carrier PC2 and a thirdcarrier PC3, respectively. Namely, each carrier PC1, PC2 or PC3pivotally supports the corresponding pinion P1, P2 or P3.

An input shaft IN is always connected (or engaged) with first carrierPC1 (i.e., constantly rotates together with first carrier PC1). Thirdring gear R3 is always connected with second carrier PC2 to form a firstrotating member M1. Second ring gear R2 is always connected with thirdcarrier PC3 to form a second rotating member M2. An output shaft OUT isalways connected with first sun gear S1. Third sun gear S3 is alwayslocked relative to a transmission case 1 (i.e., constantly fastened totransmission case 1).

The automatic transmission further includes five clutches, i.e., firstto fifth friction elements A, B, C, D and E. The first friction elementA is provided between third carrier PC3 and first sun gear S1, and isadapted to selectively connect (engage) third carrier PC3 with first sungear S1. The second friction element B is provided between first carrierPC1 and second sun gear S2, and is adapted to selectively connect firstcarrier PC1 with second sun gear S2. The third friction element C isprovided between first ring gear R1 and second sun gear S2, and isadapted to selectively connect first ring gear R1 with second sun gearS2. The fourth friction element D is provided between first carrier PC1and first rotating member M1, and is adapted to selectively connectfirst carrier PC1 with first rotating member M1. The fifth frictionelement E is provided between first ring gear R1 and second rotatingmember M2, and is adapted to selectively connect first ring gear R1 withsecond rotating member M2. In the fourth embodiment, the three planetarygear sets are disposed in order as third planetary gear set PG3→secondplanetary gear set PG2→first planetary gear set PG1 from the side ofplanetary gear set connected with input shaft IN and output shaft OUT.

Output shaft OUT is provided with an output gear or the like to transmitrotational driving force through a differential gear and a drive shaftto a drive wheel, which are not shown. In the case of the fourthembodiment, since output shaft OUT is not obstructed by the other memberor the like, the automatic transmission is applicable to both of the FFvehicle and FR vehicle.

The relations in engagements of the friction elements under respectivespeed-ratios (i.e., respective steps for shift) will be explained belowreferring to the engagement table of FIG. 14 (these engagements forrespective speed-ratios are attained by a shift control section ormeans). In the table of FIG. 14, the sign ◯ represents the engagement,and the blank represents the disengagement.

At first, the states at the time of forward running will now beexplained. A first-speed (first speed-ratio) is achieved by engagingfirst friction element A and second friction element B. A second-speedis achieved by engaging first friction element A and third frictionelement C. A third-speed is achieved by engaging first friction elementA and fourth friction element D. A fourth-speed is achieved by engagingfirst friction element A and fifth friction element E. A fifth-speed isachieved by engaging fourth friction element D and fifth frictionelement E. A sixth-speed is achieved by engaging second friction elementB and fifth friction element E. A seventh-speed is achieved by engagingthird friction element C and fifth friction element E. Next, the stateat the time of reverse running is now explained. A reverse-speed isachieved by engaging third friction element C and fourth frictionelement D.

Next, a concrete example of the reduction gear ratios according to thefourth embodiment will now be explained referring to FIG. 14. Thefollowing explanations are given in the case where a gear ratioρ1=ZS1/ZR1 of first planetary gear set PG1 is equal to 0.60 (i.e.,ρ1=ZS1/ZR1=0.60), a gear ratio ρ2=ZS2/ZR2 of second planetary gear setPG2 is equal to 0.40 (i.e., ρ2=ZS2/ZR2=0.40), and a gear ratioρ3=ZS3/ZR3 of third planetary gear set PG3 is equal to 0.50 (i.e.,ρ3=ZS3/ZR3=0.50). Where, each of ZS1, ZS2, ZS3, ZR1, ZR2 and ZR3represents the number of teeth of the corresponding gear.

A reduction gear ratio i1 of the first-speed in the forward running isexpressed by a formula: i1=1+ρ3+ρ3/ρ2. By assigning the concretenumerical values to this formula, reduction gear ratio i1 of the forwardfirst-speed is calculated as i1=2.750. The inverse of reduction gearratio i1 is equal to 0.371.

A reduction gear ratio i2 of the second-speed in the forward running isexpressed by a formula: i2=1+ρ3(1+ρ2)/(ρ2(1+ρ1)). By assigning theconcrete numerical values to this formula, reduction gear ratio i2 ofthe forward second-speed is calculated as i2=2.094. The inverse ofreduction gear ratio i2 is equal to 0.478.

A reduction gear ratio i3 of the third-speed in the forward running isexpressed by a formula: i3=1+ρ3. By assigning the concrete numericalvalues to this formula, reduction gear ratio i3 of the forwardthird-speed is calculated as i3=1.500. The inverse of reduction gearratio i3 is equal to 0.667.

A reduction gear ratio i4 of the fourth-speed in the forward running isexpressed by a formula: i4=1.0. Without assigning the concrete numericalvalues to this formula, reduction gear ratio i4 of the forwardfourth-speed is equal to 1.000. The inverse of reduction gear ratio i4is equal to 1.000.

A reduction gear ratio i5 of the fifth-speed in the forward running isexpressed by a formula: i5=ρ1(1+ρ3)/(ρ1+ρ3+ρ1 ρ3). By assigning theconcrete numerical values to this formula, reduction gear ratio i5 ofthe forward fifth-speed is calculated as i5=0.643. The inverse ofreduction gear ratio i5 is equal to 1.555.

A reduction gear ratio i6 of the sixth-speed in the forward running isexpressed by a formula: i6=ρ1(ρ2+ρ3+ρ2 ρ3)/(ρ1 ρ2+ρ3(1+ρ1)(1+ρ2)). Byassigning the concrete numerical values to this formula, reduction gearratio i6 of the forward sixth-speed is calculated as i6=0.485. Theinverse of reduction gear ratio i6 is equal to 2.062.

A reduction gear ratio i7 of the seventh-speed in the forward running isexpressed by a formula: i7=ρ1/(1+ρ1). By assigning the concretenumerical values to this formula, reduction gear ratio i7 of the forwardseventh-speed is calculated as i7=0.375. The inverse of reduction gearratio i7 is equal to 2.667.

A reduction gear ratio iR of the reverse-speed is expressed by aformula: iR=ρ1ρ2(1+ρ3)/(ρ1ρ2+ρ1ρ2ρ3−ρ3). By assigning the concretenumerical values to this formula, reduction gear ratio iR of thereverse-speed is calculated as iR=−2.571. The inverse of reduction gearratio iR is equal to −0.389.

Effects According to the Fourth Embodiment

{circle around (1)} Effects by Virtue of Structural Skeleton as a Whole

In the fourth embodiment, the automatic transmission capable ofattaining the seven speeds of forward running and the one speed ofreverse running can be obtained with proper reduction gear ratios beingensured; although the automatic transmission is constructed by a limitednumber of simple constructional elements, namely, the three sets ofsimple planet gears (three single-pinion-type planetary gear sets) andthe five friction elements.

{circle around (2)} Effects by Virtue of the Usage of Three SimplePlanetary Gear Sets

Because of the usage of the three sets of simple planet gears (the usageof three single-pinion-type planetary gear sets), a gear noise and atransfer efficiency can be improved as compared to the case where doublepinions (double-pinion-type planetary gear set) are used. Further,because a diameter of the pinion does not need to be reduced in thisembodiment, a gear durability can be enhanced.

{circle around (3)} Effects Based on Gear Ratios

All of the gear ratios ρ1, ρ2 and ρ3 in the respective planetary gearsets are close to a center value 0.5. Accordingly, a possible rangewhich is obtained by freely setting the three gear ratios is wide sothat a degree of freedom in reduction gear ratio can become higher.

{circle around (4)} Effects Based on a Ratio Coverage in the ForwardRunning

A ratio coverage (gear-ratio width) of the forward running is defined bydividing the reduction gear ratio of the lowest-speed (step) by thereduction gear ratio of the highest-speed, i.e., the reduction gearratio of the lowest-speed/the reduction gear ratio of the highest-speed.A compatibility between an accelerating performance at the time ofvehicle start and a fuel economy at the time of high speed cruise ofvehicle becomes better, and also a degree of freedom to set the gearratio value in respective forward speed-ratios becomes higher; as thevalue of ratio coverage becomes greater. As concrete numerical values inthe fourth embodiment, the reduction gear ratio of the forwardfirst-speed is equal to 2.750 and the reduction gear ratio of theforward seventh-speed is equal to 0.375. In this embodiment, the ratiocoverage from first-speed to seventh-speed is equal to 7.33, and hence asufficient ratio coverage can be ensured. Therefore, for example, theautomatic transmission according to the fourth embodiment is useful alsoas a transmission for a vehicle equipped with a diesel engine as itspower source, although a width of rotational speed (number ofrevolutions) of diesel engine is narrower than that of a gasoline engineand a torque of diesel engine is higher than that of a gasoline enginehaving the same engine displacement.

Moreover, in the case where the gear ratio value of the low-speed sideis great relative to the ratio coverage, a torque transmitted to a finalgear becomes relatively great. Hence, this case requires a sufficientstrength of the automatic transmission or propeller shaft, so that thewhole of vehicle body is upsized. That is, it is preferable that thegear ratio value (value of speed ratio) of the lowest-speed is not sogreat under the same condition of ratio coverage. In an automatictransmission shown by FIG. 23 in the document of U.S. Pat. No.6,648,791, a gear ratio of the highest-speed (highest step for shift) isequal to 1. Hence in this technique, when trying to enlarge the ratiocoverage, a gear ratio (value) of the lowest-speed needs to be enlarged,so that the upsizing of the automatic transmission and the propellershaft is caused. On the other hand, in the automatic transmissionaccording to the fourth embodiment of the present invention, asufficient ratio coverage can be ensured without the necessity ofenlarging the gear ratio of the lowest-speed so much.

{circle around (5)} Effects Based on 1-R Ratio

A value of 1-R ratio (Reverse-speed/First-speed) is a value near 1,concretely, equal to 0.935. Accordingly, an acceleration feel of vehiclerelative to a depressing adjustment of an accelerator pedal does notgreatly vary between the forward running and the reverse running.Therefore, the problem that the drivability worsens can be avoided.

{circle around (6)} Effects Based on the Number of Changeovers Among theFriction Elements at the Time of Shift

(i) If one or more friction element is released and two or more frictionelements are engaged at the time of shift, or if two or more frictionelements are released and one or more friction element is engaged at thetime of shift; a torque control and a control for the engaging andreleasing timings of friction elements become complicated. Hence, from aviewpoint of avoidance of the complication of shift control, it isfavorable that one friction element is released and one friction elementis engaged at the time of shift. That is, it is favorable that aso-called double-changeover is avoided. In the fourth embodiment, theshifts among the forward first-speed to the forward fourth-speed areperformed under the condition where first friction element A ismaintained in engaged state. Moreover, the shifts among the forwardfourth-speed to the forward seventh-speed are performed under thecondition where fifth friction element E is maintained in engaged state.Namely, each shift between adjacent two speed-ratios (gear steps) amongthe forward first-speed to seventh-speed can be achieved by releasingone friction element and by engaging one friction element. Accordingly,each of all the shifts between adjacent two speed-ratios of forwardrunning is performed by means of only the changeover from one frictionelement to the other one friction element. Therefore, the control duringthe shift can be prevented from being complicated.

(ii) As mentioned in the above (i), all the shifts between adjacent twospeed-ratios can be achieved by the changeover shift which releases onefriction element and engages one friction element. Moreover, similarly,each one-ratio-skip shift (e.g., first-speed→third-speed) among theforward first- to seventh-speeds can also be achieved by releasing onefriction element and by engaging the other one friction element.Accordingly, a controllability thereof can be enhanced.

{circle around (7)} Effects Based on Layout

(i) In the automatic transmission according to the fourth embodiment, ona radially outer side of the three planetary gear sets, connectingmembers are disposed so as not to become in a three-layered form, asshown in the skeleton diagram of FIG. 13. Accordingly, it is difficultto cause the retention of a lubricating oil (disrupted flow oflubricating oil), so that the fuel economy can be improved by reducingthe frictions.

(ii) Moreover, the rotating member passing on the outer peripheral sideof the planetary gear sets is formed in a single-layered structure, asshown by the skeleton diagram of FIG. 13. Generally in the automatictransmission, lubricating oil is always supplied to respective rotatingelements such as gears and bearings (not shown) for the purpose ofcooling, lubrication and the like. This lubricating oil is generallysupplied from a shaft-center side of the transmission by means ofcentrifugal force. At this time, if an efficiency of discharge oflubricating oil becomes worsened on the outer peripheral side of theplanetary gear sets, oil temperature rises so that a durability offriction elements, bearings and the like is reduced. Since the rotatingmember passing on the outer peripheral side of the planetary gear setsforms the single-layered structure in the fourth embodiment as mentionedabove, the discharging efficiency of lubricating oil is not worsened sothat the temperature rise is suppressed to improve the durability.

(iii) The automatic transmission according to the fourth embodiment canbe designed to allow torque to be inputted to one side of the planetarygear sets and then to be outputted from another side of the planetarygear sets. Accordingly, the automatic transmission according to thefourth embodiment is applicable to both of the front drive vehicle andrear drive vehicle, namely can be widely applied.

This application is based on a prior Japanese Patent Application No.2007-309539 filed on Nov. 29, 2007. The entire contents of this JapanesePatent Application are hereby incorporated by reference.

Although the invention has been described above with reference tocertain embodiments of the invention, the invention is not limited tothe embodiments described above. Modifications and variations of theembodiments described above will occur to those skilled in the art inlight of the above teachings. The scope of the invention is defined withreference to the following claims.

1. An automatic transmission comprising: a first planetary gear setincluding a first sun gear, a first pinion engaged with the first sungear, a first carrier supporting the first pinion, and a first ring gearengaged with the first pinion and constantly locked; a second planetarygear set including a second sun gear constantly locked, a second pinionengaged with the second sun gear, a second carrier supporting the secondpinion, and a second ring gear engaged with the second pinion, andconnected with the first carrier to define a rotating member; a thirdplanetary gear set including a third sun gear, a third pinion engagedwith the third sun gear, a third carrier supporting the third pinion,and a third ring gear engaged with the third pinion; an input shaftconstantly connected with the third carrier; an output shaft constantlyconnected with third ring gear; and five friction elements including afirst friction element adapted to selectively connect the second carrierwith the third ring gear, a second friction element adapted toselectively connect the first sun gear with the third carrier, a thirdfriction element adapted to selectively connect the first sun gear withthe third sun gear, a fourth friction element adapted to selectivelyconnect the rotating member with the third carrier, and a fifth frictionelement adapted to selectively connect the rotating member with thethird sun gear, the automatic transmission being adapted to achieve atleast seven forward speed-ratios and one reverse speed-ratio, each ofthe at least seven forward speed-ratios and one reverse speed-ratiobeing achieved by an engaged state of two friction elements selectedfrom the five friction elements.
 2. The automatic transmission asclaimed in claim 1, wherein the seven forward speed-ratios are achievedby concurrent engagements of the first friction element and the secondfriction element, concurrent engagements of the first friction elementand the third friction element, concurrent engagements of the firstfriction element and the fourth friction element, concurrent engagementsof the first friction element and the fifth friction element, concurrentengagements of the fourth friction element and the fifth frictionelement, concurrent engagements of the second friction element and thefifth friction element, and concurrent engagements of the third frictionelement and the fifth friction element.
 3. The automatic transmission asclaimed in claim 2, wherein the one reverse speed-ratio is achieved byconcurrent engagements of the third friction element and the fourthfriction element.
 4. The automatic transmission as claimed in claim 1,wherein each of all the shifts between adjacent two speed-ratios amongthe at least seven forward speed-ratios is performed by releasing onefriction element of the at least five friction elements and by engagingthe other one friction element of the at least five friction elements.